Tensioner with increased damping

ABSTRACT

In an aspect, a tensioner for an endless drive member, comprising a shaft and base that are mountable to be stationary relative to an engine, a tensioner arm that is pivotable relative to the shaft about a tensioner arm axis, a pulley on the tensioner arm rotatable about a pulley axis that is offset from the tensioner arm axis, and that is engageable with an endless drive member, a tensioner spring that is positioned to urge the tensioner arm towards a free arm position, a damping element that engages the tensioner arm and that is engaged by a plurality of axially spaced segments of the tensioner spring.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to U.S. Provisional Patent ApplicationNo. 61/716,894 filed Oct. 22, 2012 the contents of which areincorporated herein in their entirety.

FIELD

This disclosure relates to tensioners and in particular tensioners thatoperate to tension synchronous endless drive members such as a timingbelt on an engine.

BACKGROUND

Tensioners are known devices for maintaining tension in belts (e.g.timing belts) or other endless drive members that are driven by anengine and that are used to drive certain components, such as camshafts.A tensioner typically includes a base that mounts to the engine, atensioner arm that is pivotable with respect to the base about a pivotaxis, a pulley that is mounted at a free end of the arm for engagementwith the belt, and a spring that acts between the base and the arm todrive the arm into the belt. The direction into the belt (i.e. thedirection in which the spring drives the arm) may be referred to as adirection towards a free arm position (i.e. towards a position that thetensioner arm would reach if no belt were present to stop it). This is adirection of lessening spring potential energy. The tensioner arm ingeneral moves in this direction as the belt tension drops. The directionaway from the belt (i.e. the direction against the biasing force of thespring) may be referred to as a direction towards a load stop position,and is a direction of increasing spring potential energy. The tensionerarm in general moves in this direction as the belt tension increases. Itis known that it is desirable to provide damping on a tensioner in orderto assist the tensioner arm in resisting being thrown off a belt duringsudden increases in belt tension which can accelerate the tensioner armsuddenly towards the load stop position. In at least some demandingapplications, however, the damping that is available from a typicalprior art tensioner is not sufficient to satisfactorily inhibit such anevent from happening. It would be desirable to provide a tensioner thathas increased damping.

SUMMARY

In an aspect, a tensioner for an endless drive member, comprising ashaft and base that are mountable to be stationary relative to anengine, a tensioner arm that is pivotable relative to the shaft about atensioner arm axis, a pulley on the tensioner arm rotatable about apulley axis that is offset from the tensioner arm axis, and that isengageable with an endless drive member, a tensioner spring that ispositioned to urge the tensioner arm towards a free arm position, adamping element that engages the tensioner arm and that is engaged by aplurality of axially spaced segments of the tensioner spring.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a sectional side view of a prior art tensioner;

FIG. 2A is a sectional side view of another prior art tensioner;

FIG. 2B is a magnified sectional side view of a portion of the tensionershown in FIG. 2;

FIG. 3 is a side view of an engine with a tensioner in accordance withan embodiment of the present invention;

FIG. 4 is a sectional side view of the tensioner shown in FIG. 3;

FIGS. 4A and 4B are exploded perspective views of a variant of thetensioner shown in FIG. 4;

FIG. 5 is a magnified sectional side view of a portion of the tensionershown in FIG. 4;

FIG. 6 is a sectional view of a ‘bottom’ or ‘proximal’ portion of thetensioner shown in FIG. 4, that contacts the engine;

FIG. 7 is a sectional view of an ‘upper’ or ‘distal’ portion of thetensioner shown in FIG. 4;

FIG. 8 is a plan view of a spring and a damping element from thetensioner shown in FIG. 4;

FIG. 8A is a perspective view of the spring from the tensioner shown inFIG. 4;

FIG. 9 is a sectional side view of the spring and the damping elementshown in FIG. 8;

FIG. 10 is a sectional side view of a tensioner in accordance withanother embodiment of the present invention;

FIG. 11 is a magnified sectional side view of a portion of the tensionershown in FIG. 10;

FIG. 12 is a graph illustrating the hub load on the pulley from thetensioners shown in FIG. 1, FIG. 2 and FIG. 4;

FIGS. 13 and 14 are sectional side and plan views of another prior arttensioner;

FIG. 15A-15G are graphs illustrating the belt tension in relation to thearm position of several different tensioner configurations;

FIG. 16 is a diagram illustrating the setup for the tensioners testedfrom which the curves in FIGS. 15A-15G were generated;

FIGS. 17-19 are plan views of the spring and the damping element, withdifferent relative angles between two ends of the spring; and

FIG. 20 is a diagram illustrating the vector sum of the forces acting onthe ends of the spring in either of FIGS. 18 and 19.

DETAILED DESCRIPTION OF EXAMPLE EMBODIMENTS

A prior art tensioner is shown at 10 in FIG. 1, and includes a dampingstructure 12 that absorbs kinetic energy from a timing drive andconverts the kinetic energy to heat through friction between componentsof the damping structure 12. The tensioner 10 includes a shaft 14, abase 15, a bushing 16, a tensioner arm 18, a pulley 20 that rotates onthe arm 18 via a bearing 21 (e.g. a ball bearing) and a tensioner spring22. The bushing 16 pivots with the tensioner arm about the shaft 14during operation of the tensioner 10 in response to changes in tensionof the endless drive member against which the pulley 20 is engaged. Theendless drive member is not shown in FIG. 1, but it will be understoodthat it may be a timing belt or the like.

One source of friction is between the shaft 14 and the oscillatingbushing 16 (which may be termed ‘shaft-bushing friction’). In someapplications the shaft-bushing friction is sufficient to control timingdrive dynamics. However, sometimes more friction is required, asinsufficient damping may lead to catastrophic failure of the belt andconsequently catastrophic failure of the engine particularly ininterference engine designs where the valves could collide with thepistons if the valve timing is incorrect. Referring to FIG. 2A, a springsupport 24 is provided. The spring support 24 may in some instances berotationally locked to the base 15 by a locking feature. Also, the axialforce of the spring 22 against the spring support 24 can at leastsomewhat prevent rotation of the spring support 24 relative to the base15 (i.e. a high frictional torque may exist between the spring support24 and the base 15 which means that the spring support 24 is not ‘fixed’to the base 15 but would not be expected to rotate under most operatingconditions. The first coil of the spring 22 engages the spring support24 and applies a force F on it radially inwardly towards the tensionerarm pivot axis A_(a). An example of the spring support is shown in FIGS.13 and 14, which are figures from U.S. Pat. No. 4,473,362, the contentsof which are hereby incorporated by reference. It will be noted thatFIGS. 13 and 14 are reproductions of figures from the aforementioned USpatent including reference numerals used in that patent. Accordingly,the reference numerals in those two figures do not relate to items inthis description. For example, item 112 as used in this description isnot related to reference numeral 112 as it appears in FIGS. 13 and 14.

The spring support 24 may be made of nylon or any other suitablematerial. The compressive force F of the spring 22 on the spring support24 urges the spring support 24 against the tensioner arm 18, whichgenerates friction as the tensioner arm 18 pivots and slides against thespring support 24 during operation of the tensioner 10.

The spring 22 (which may be a torsion spring as shown) generates atorque T. The spring 22 has first and second ends 23 and 25 which end intangs that are not shown in the sectional views in FIGS. 2A and 2B butwhich engage the base 15 and the arm 18 respectively. The first end 23is hooked to the base 15 so as to be stationary relative to the engine,and is positioned at a distance r_(c) from the spring centre. The secondend 25 moves with the tensioner arm 18. Thus along the helical length ofthe spring 22, there is progressively more and more rotationaloscillation movement starting from the stationary first end 23 along thelength of the spring 22 to the oscillating second end 25. A force F actson the first end 23 of the spring 22. This force is transmitted to thespring support 24. The friction coefficient between the engaged surfacesof the arm 18 and the spring support 24 is represented by μ_(a). Theradius of the arm 18 at the region where the arm 18 contacts the springsupport 24 is represented by r_(a) as shown in FIG. 2B. The frictionaltorque generated by the spring support 24 is: M_(a)=r_(a)*μ_(a)*F.Additionally, the compression of the spring support 24 in turncompresses the arm 18 by some amount, which in turn causes additionalshaft-bushing friction. This additional frictional torque will beignored for the purposes of this description. The shaft-bushingfrictional torque (ignoring compressive effects from the spring support24) may be calculated as: M_(b)=r_(b)*μ_(b)*H₁, where r_(b) is thebushing radius, μ_(b) is the friction coefficient between the engagedsurfaces of the bushing 16 and the shaft 14, and H₁ is the hubloadvector. In an example, where the tensioner 10 has a 3 mm arm (i.e. theoffset between the pulley axis shown as A_(p) and the tensioner arm axisA_(a) is 3 mm), and assuming an angle between the tensioner arm 18 andthe hubload vector of 90°, the frictional torque generated by the springsupport 24 is approximately 100%*M_(a)/M_(b)=54% of the frictionaltorque generated by bushing (where r_(a)=12.5 mm, R=3 mm, μ_(a)=0.2,r_(c)=20 mm, r_(b)=10 mm, μ_(b)=0.07). Thus the frictional torqueprovided by the spring support 24 may be substantial, compared to thefrictional torque provided by the engagement between the shaft 14 andbushing 16.

It would be advantageous to provide other sources of friction, asidefrom those described above in at least some situations. For example, inbelt-in-oil applications the friction generated between the shaft 14 andbushing 16 is reduced due to lubrication caused by the presence of theoil. Thus an additional source or alternative source of friction isdesirable. Also, in many applications it would be advantageous to beable to use a relatively longer tensioner arm (i.e. where R>3 mm). Alonger tensioner arm may have better/larger take-up than a shorter arm,where ‘take-up’ is the amount of belt length the tensioner cancompensate per one degree of rotation of the tensioner arm.Additionally, a longer arm permits more stable tension control. A longerarm may permit relatively easy pull-the-pin installation of thetensioner as compared to some tensioners which are complicated toinstall due to their use of an installation eccentric, which is anoffset between the center of the shaft 14 and a pivot axis of the shaft14 that is used to adjust the position of the shaft 14 duringinstallation of the tensioner. Once the shaft 14 is correctlypositioned, it is fixedly captured in its current position using a boltor the like so that it does not pivot. Installation of such tensioners,however, can be complicated as noted above.

Long arm tensioners of the prior art, however, can sometimes generatetoo little frictional torque at the shaft-bushing interface precludingtheir use in some situations. In order to keep the bearings in thetensioner small (to keep costs down), an increase of the arm length canlead to a reduced shaft diameter, which in turn results in lessfrictional torque at the shaft-bushing interface. In an example, for atensioner with a 30 mm inner diameter ball bearing, the ratio of thefrictional torque to produced torque may be determined as follows: A 3mm arm may be packaged with a 20 mm diameter shaft. It may have a 15 mmbearing radius−3 mm (arm eccentric)−1 mm (aluminum arm wall)−1 mm(bushing thickness)=10 mm (shaft radius). A 5 mm arm may only bepackaged in the same bearing with 16 mm diameter shaft using a similarcalculation as above. For a 3 mm arm, the damping ratio would be 10 mm(shaft radius)*0.1 (friction coefficient)*F (load)/3 mm (armeccentric)*F=0.33. For a 5 mm arm, the damping ratio would be8*0.1/5=0.16. With such a low damping ratio, the tensioner with a 5 mmarm would not be able to control timing drive dynamics in some cases. Adamping ratio in the range of about 0.3 to about 0.4 may be suitable insome applications.

A tensioner 100 as shown mounted to an engine 101 in FIG. 3, whichprovides additional damping as compared to the tensioner shown in FIGS.2A and 2B. The tensioner 100 acts on a timing belt 103 that transfersrotational power from a crankshaft 104 to a pair of camshafts 105 a and105 b. The additional damping provided by the tensioner 100 is providedvia a spring support as shown at 124 in FIG. 4. As a result, thetensioner 100 may have a tensioner arm length of more than 3 mm (e.g. 5mm) in some instances. The tensioner 100 has a shaft 114 that may besimilar to the shaft 14, a base 115 that is staked to one end of theshaft 114, a bushing 116 that may be similar to bushing 16, a tensionerarm 118 that may be similar to the tensioner arm 18, a pulley 120 thatmay be similar to the pulley 20, a bearing 121 that may be similar tothe bearing 21, and a tensioner spring 122 that may be similar to thespring 22.

The tensioner arm 118 is pivotable about a tensioner arm pivot axis Aashown in FIG. 4. The pulley 120 is rotatable about a pulley axis Ap,which is offset from the tensioner arm pivot axis Aa, wherein the amountof offset is the length of the tensioner arm.

A retaining washer 135 is staked to the other end of the shaft 114 tohold selected components together. A polymeric bushing plate 137 isprovided between the retaining washer 135 and the tensioner arm 118 toprevent metal-to-metal contact therebetween. The bushing 116 and thedamping element 124 may together be generally referred to as a dampingsystem 112.

In the embodiment shown in FIG. 4 an installation eccentric 139 isprovided in an aperture 141 in the shaft 114, which permits adjustmentof position of the tensioner arm 118 during installation of thetensioner 100 on an engine. A fastener shown at 119 in FIG. 3, passesthrough the aperture 141 (FIG. 4) but is offset from the center ofaperture 141 by the installation eccentric 139 to mount the tensioner100 to the engine. However, in a preferred variant shown in FIGS. 4B and4B, there is no installation eccentric, and a longer tensioner armlength 118 can be provided (e.g. 5 mm as opposed to 3 mm). In thisvariant, a fastener (not shown) passes through the aperture 141 (and iscentered therein), to mount the tensioner 100 to the engine. The variantshown in FIGS. 4A and 4B may otherwise be similar to the embodimentshown in FIG. 4.

FIG. 6 is a sectional view of a ‘bottom’ or ‘proximal’ portion of thetensioner 100 that contacts the engine. FIG. 7 is a sectional view of an‘upper’ or ‘distal’ portion of the tensioner 100. FIG. 8 is a plan viewof the spring 122 and the damping element 124. FIG. 8A is a perspectiveview of the spring 122 alone. FIG. 9 is a sectional side view of thespring 122 and the damping element 124.

As shown in FIGS. 8 and 8A, the tensioner spring 122 has a first end 123and a second end 125, each of which ends in a tang. As shown in FIGS. 4Aand 6, the tang at the first end 123 engages the base 115 (moreparticularly it engages a slot 111 in the base 115) so as to anchor thefirst end 123 of the spring 122. Additionally, as shown in FIG. 7 thetang at the first end 123 passes through a slot 117 in the dampingelement 124, which rotationally fixes the damping element 124 to thebase 115, while still permitting the damping element 124 to slide asneeded to engage the tensioner arm 118. As shown in FIGS. 4B and 7 thetang at the second end 125 engages the tensioner arm 118 (moreparticularly it engages a slot 113 in the tensioner arm 118) so as toapply a biasing force urging the tensioner arm 118 into the belt 103.Referring to FIG. 8A, the spring 122 may be a helical torsion springthat includes more than one coil wherein a coil is defined as a segmentof the spring 122 that extends through 360 degrees. In this embodimentthe spring 122 has 2.5 coils including a first end coil 129 a, a secondend coil 129 b, and a 180 degree segment shown at 129 c between thefirst and second end coils 129 a and 129 b. Delimiters between the coils129 a, 129 b and the segment 129 c are shown at 131.

Referring to FIGS. 8, 8A and 9, the spring support 124 may be similar tothe spring support 24, however, in the tensioner 100, the spring support124 is configured such that a first segment 127 a in the first end coil129 a and a second segment 127 b in the second end coil 129 b bothengage the spring support 124. The segments 127 a and 127 b are shown ina perspective view in FIG. 8A. In FIG. 6, only a portion of the spring122 is shown so that the segment 127 a is not obstructed. In FIG. 7, adifferent portion of the spring 122 is shown so that the segment 127 bcan be seen unobstructed. As can be seen in FIGS. 6, 7 and 8A,delimiters shown at 133 show the extents of the segments 127 a and 127 b(i.e. they show the ends of the portions of the spring 122 that contactthe spring support 124). As can be seen, the first segment 127 a and thesecond segment 127 b are axially offset or spaced from one another, andare in the first end coil 129 a and second end 129 b, respectively.Also, as can be seen in FIG. 8A, the first and second segments 127 a and127 b are generally aligned axially.

As shown in FIG. 8, the overall force exerted on the spring support 124by the spring 122 is F at the first end 123 and F at the second end 125.As a result, the overall frictional torque (and therefore damping)generated by the spring support 124 and the arm 118 is greater than(i.e. approximately double) that produced in the embodiment shown inFIGS. 2A and 2B (and the embodiment shown in FIG. 1). This is based onthe assumption that the force F that is exerted at the two ends 123 and125 of the spring 122 are approximately in the same direction, which istrue when the spring ends 123 and 125 are approximately 180 degreesapart angularly about the spring axis, shown at As (FIG. 8). Theirrelative positions are shown in an example in FIG. 8. Their ranges ofrelative positions during use over a range of belt tensions are shown inFIGS. 17-19. When they are 180 degrees apart, the forces are aligned andare purely additive. When they are at some other angle relative to eachother, the forces are not purely additive and vector components of theforces must be considered to determine the overall force exerted on thespring support 124.

As noted above the second spring end 125 oscillates together with thetensioner arm 118. Thus, there is friction generated between secondspring end 125 and the stationary spring support 124, and consequentlythere is the potential for wear on the spring support 124. In abelt-in-oil application this wear may be acceptable. In applicationswhere the wire used for the spring 122 has a square cross-sectionalshape, the pressure of the spring 122 on the spring support 124 is lowerthan the pressure exerted by a spring 122 round cross-section (sincethere is more contact area on the square cross-section spring). Thus inembodiments where a square (or rectangular) cross-section spring 122 isused, such as is shown in FIG. 4, the wear may be acceptable. The wearmay also be acceptable even in embodiments that include a roundcross-section spring 122.

Based on the above, it can be seen that, as compared to the embodimentshown in FIGS. 2A and 2B, the frictional torque generated using theembodiment shown in FIG. 4 is as follows: approximately twice thefrictional torque generated by the spring support 24, a limitedly higherfrictional torque that is generated between the shaft 14 and bushing 16,and additional friction generated between the second end 125 of thespring 122 and the spring support 124, which has no analogousarrangement in the embodiment in FIGS. 2A and 2B. The additional dampingprovided by the embodiment shown in FIG. 4 may permit the use of alonger arm tensioner than is possible with the embodiment shown in FIGS.2A and 2B. It may also permit the tensioner 100 to be used in abelt-in-oil application, in either a short-arm or long-armconfiguration. The spring support 124 may be referred to as a dampingelement or a damper due to its increase role in the damping.

It will be noted that, while the spring 122 constricts during operationand applies a compressive force on the damping element 124 it may pushoil away (in a helical direction—along the length of the spring coils)from the contact area between the spring 122 and the damping element124. As a result, the presence of oil may not cause a large reduction infriction between the spring 122 and the damping element 124. It willalso be noted that, while the spring 122 may be made from a spring wirehaving a square or rectangular cross-section, the wire may twist by someangle during operation and thus may engage the damping element at somepoints along a corner of the cross-sectional shape and not along a flatface of the cross-sectional shape. This will reduce by some amountlubricating effects of any oil that is present that would reduce thefriction between the spring 122 and the damping element 124.

As shown in FIGS. 10 and 11, instead of only engaging the dampingelement 124 with segments 127 a and 127 b in the first and second endcoils 129 a and 129 b of the spring 122 as shown in FIGS. 4 and 5, it isalternatively possible to provide an embodiment where at least oneadditional spring segment engages the damping element 124. For example,in the embodiment shown in FIGS. 10 and 11, the segment shown at 127 c,which is axially between the first and segments 127 a and 127 b, alsoengages the damping element 124. The overall force exerted by the spring122 remains as being 2F, (based on the force F that is applied to thespring 122 at both ends 123 and 125 as shown in FIGS. 8 and 9), butbecause there are three spring segments that contact the damping element124, the force exerted on the damping element 124 by each of the threespring segments 127 a, 127 b and 127 c, is 2F/3, as shown in FIGS. 10and 11. As a result, there is a reduced pressure applied by each coil onthe damping element 124 as compared to the embodiment shown in FIGS. 4and 5. Consequently there may be less wear on the damping element 124.If more than three axially spaced segments are in contact with thedamping element 124, the force applied by each coil may be reducedfurther, thereby reducing the pressure on the damping element whilemaintaining the overall force (i.e. 2F).

As noted above, all along the helical length of the spring 122, themagnitude of the oscillatory movement increases progressively along thehelical length of the spring 122 from the stationary first end 123 tothe second end 125 which oscillates with the arm 118. Thus each segment127 a, 127 c, 127 b has progressively more sliding movement with thedamping element 124. While the first segment 127 a has some non-zeroamount of sliding movement with the damping element 124 it is relativelysmall and may be ignored for its impact on the overall damping providedby the tensioner 100. In an example if there are three segments 127 a,127 b, and 127 c and the tensioner arm 118 (and therefore the second end125) oscillates with an amplitude of ±6 degrees, then the third segment127 c would oscillate with an amplitude of about ±3 degrees, and thefirst segment 127 a would substantially not oscillate). In theembodiment shown in FIGS. 10 and 11, the frictional torque provided bythe damping element 124 may be about 2.5 times the frictional torqueprovided by the spring support 24 of the tensioner 10 shown in FIGS. 2Aand 2B. This increase in frictional torque (and therefore damping) isprovided at essentially no added cost or complexity and without addingnew components. The performance of the tensioner 100 relative to thetensioner 10 is shown in FIG. 12, which shows hysteresis curves for thetensioner 100 at 180, the tensioner 10 with a spring support at 182 andthe tensioner 10 with no damping element (as shown in FIG. 1) at 184.

In applications where the friction surfaces will be exposed to oil,features may be provided to assist in removing oil from them so as toreduce the risk of sudden drops in friction and damping that can occurfrom the presence of oil. Slits (which may be referred to as channels orgrooves) shown at 186 in FIG. 9 in the damping element 124 are designedto provide oil reservoirs or transport channels to help to transfer oilout from contacting surfaces between the damping element 124 and the arm118 (not shown in FIG. 9). The channels 186 may be provided with sharpedges which can scrape oil from the surface of the arm 118 to reduce therisk of development of an oil film between the contacting surfaces ofthe damping element 124 and the arm 118. The presence of the channels186 reduces the overall contact area between the damping element 124 andthe arm 118, which increases the surface pressure between them. The sizeand/or number of the channels 186 can be selected to provide a selectedsurface pressure that may be high enough to squeeze oil out from betweenthe contacting surfaces.

Another feature that may be provided on the damping element 124 may bereductions in wall thickness 188, which may be referred to as flexjoints 188. The flex joints 188 increase the flexibility of the dampingelement wall (shown in FIG. 9 at 190) which increases the contact areabetween the wall 190 and the tensioner arm 118 which in turn makes forless wear and more stable friction between them. These flex joints 188can be provided any suitable way. For example, the flex joints 188 mayextend axially along the axial length of the damping element wall 190.They may, in some embodiments, be provided on the face of the wall 190that faces the arm 118. In the embodiment shown in FIGS. 8 and 9, it canbe seen that the flex joints 188 are formed by slots that pass throughthe entirety of the wall thickness and which extend substantially alongthe entire axial length of the damping element 124 and which are closedat a first end 188 a and open at a second end 188 b. As can be seen inFIG. 8 in particular, these slots separate a portion of the dampingelement 124 into segments shown at 191.

In other embodiments, the wall thickness at these flex joints 188 may beabout half of the wall thickness elsewhere, or it may be a differentnon-zero fraction of the wall thickness away from the flex joints 188.

The graphs in FIGS. 15A-15G illustrate tension control provided by thetensioner 100. There are three curves shown in each graph. The top curverepresents high tension (i.e. the tensioner 100 is pushed out by thebelt shown at 192 in FIG. 16); the bottom curve is low tension (i.e. thetensioner 100 follows a slack belt 192) and the center curve is themathematical average of two extreme forces for each tensioner armposition. The graphs illustrate the performance of various tensionerconfigurations (e.g. different lengths of tensioner arm, use of springsupport 24, use of damping element 124, no damping element). The graphsin FIGS. 15A-15C represent a tensioner with a 3 mm arm length with nospring support, a tensioner with a 3 mm arm length with a spring supportsimilar to spring support 24, and a tensioner with a 3 mm arm lengthwith a damper similar to damper 124 respectively. The graphs shown inFIGS. 15D-15F represent a tensioner with a 5 mm arm length with nospring support, a tensioner with a 5 mm arm length with a spring supportsimilar to spring support 24, and a tensioner with a 5 mm arm lengthwith a damper similar to damper 124 respectively. The graph in FIG. 15Gshows a 5 mm arm tensioner with a damper similar to damper 124 and extratravel to allow for installation of a pull pin (100 degrees travel vs 62degrees travel for a standard installation method). The increased travelcreates the challenge to control tension in a stable manner. Tensioncurves from the graphs in FIGS. 15A-15G are proportional to tensionerhysteresis curves at each angular position of the tensioner arm. Thedistance between the top and bottom curves on each graph is proportionalto the tensioner damping, such that a wider distance means that moredamping is provided. The graphs are prepared assuming that damper/springcontact points are always vertically aligned, for simplicity. Thedamping element 124 reduces its influence when forces are not aligned(illustrated in FIGS. 18-19) which helps to stabilize tension control atthe extreme positions (i.e. at the ends of its range of travel). Withreduced damping, the curve representing the maximum tension force is notas steep and the distance between the maximum and minimum tensions doesnot grow as much as is shown on the graphs close to the end of thetensioner travel. This reduction in the distance between the maximum andminimum tensions at the extremes of travel is advantageous fortensioners that have pull pins that are removed after installation,which require extra travel to compensate for engine and belt buildtolerances. In some applications, the tensioner 100 may be configured tosubstantially align the damping element forces when the tensioner 100 isclose to the center of its travel as shown in FIG. 17. The alignment ofthe forces may be tailored to address a particular applicationrequirement. For example, the tensioner 100 may be configured to haveincreased resistance as it approaches its load stop position so as toreduce the likelihood of actual contact with a limit surface on thetensioner that defines the load stop position.

The graphs in FIGS. 15A-15G have the same vertical scale so as tofacilitate visually comparing the distance between the tension curves todraw conclusions as to which tensioner has more damping and will be morestable in use on an engine. It will be noted that:

-   -   the 5 mm arm tensioners have more stable tension control (i.e.        the curves are relatively flat) but less damping (i.e. there is        a smaller distance between max and min curves);    -   the 5 mm arm tensioner with the damper has more damping then the        3 mm tensioner with no damper (and will therefore oscillate less        on the engine); and    -   the tension characteristics of the 5 mm arm pull-the-pin        installation tensioner are similar (i.e. similarly parabolic) to        those of a 3 mm arm tensioner that has an installation eccentric        and its associated complicated installation procedure. Providing        a pull-the-pin installation feature requires more arm travel to        compensate for engine and belt dimensional tolerances. It can be        difficult to design a pull-the-pin type tensioner with a 3 mm        arm length due to arm “over center” condition, whereby the        tensioner arm locks up due to the geometry of the forces acting        on it.

The tensioner 100 can provide good damping even in the presence of oilfacilitates its use in a belt-in-oil arrangement, as a replacement for atiming chain design on an engine. For greater certainty however, it willbe noted that the tensioner 100 may be advantageous in applicationswhere no oil is present.

While the damping element 124 has been described as being made fromnylon, other materials may be used, such as nylon with a Teflon™ coatingon the inner surface (i.e. the surface that contacts the arm 118) or onthe outer surface (i.e. the surface that contacts the spring 122) or onboth the inner and outer surfaces. The spring 122 could be coated with alow friction material if desired so as to reduce wear that might occuron the damping element 124. Materials and coatings may be selected sothat damping and wear characteristics may be as desired for a particularapplication.

Providing the tensioner 100 which has a plurality of axially spacedsegments of the spring 122 in contact with the damping element 124permits an improvement in damping which can facilitate the use of thetensioner 100 in a belt-in-oil application and/or the use of a longerarm tensioner than is possible using some damping structures of theprior art.

While a spring 122 having 2.5 coils in helical length is shown in thefigures, it will be understood that the spring 122 could have fewer ormore than 2.5 coils. For example, the spring 122 could have 1.5 coilsand still have ends that are 180 degrees apart angularly, and wouldstill have two segments that engage the damping element 124. In anotherexample, the spring 122 could have 1.25 coils of helical length andwould have ends that are 90 degrees apart angularly, while still havingtwo segments that engage the damping element, although the forcesapplied at the spring ends would add in a vector sum of about 1.4F insuch an instance. Other spring lengths are possible, such that three,four or any suitable number of segments of the spring 122 would engage asuitably lengthened version of the damping element 124.

The above-described embodiments are intended to be examples only, andalterations and modifications may be carried out to those embodiments bythose of skill in the art.

The invention claimed is:
 1. A tensioner for an endless drive member,comprising: a shaft and a base that are mountable to be stationaryrelative to an engine; a tensioner arm that is pivotable relative to theshaft about a tensioner arm axis; a pulley on the tensioner armrotatable about a pulley axis that is offset from the tensioner armaxis, wherein the pulley is engageable with the endless drive member; atensioner spring having a first end on a first end coil that ispositioned to exert a force on the shaft and the base and a second endon a second end coil that is positioned to exert a force on thetensioner arm to urge the tensioner arm towards a free arm position; anda damping element that engages the tensioner arm and that is engaged bya plurality of axially spaced segments of the tensioner spring,including a segment that is on the second end coil having the secondend, wherein the tensioner is in a belt-in-oil arrangement.
 2. Atensioner as claimed in claim 1, wherein the plurality of axially spacedsegments are positioned on two end coils of the tensioner spring.
 3. Atensioner as claimed in claim 2, wherein the plurality of axially spacedsegments includes at least one segment of the tensioner springpositioned between the two end coils.
 4. A tensioner as claimed in claim1, wherein the tensioner spring is generally helical and has a first endand a second end, wherein torsional forces are transmitted into thespring via the first and second ends, and wherein the first and secondends are angularly 180 degrees apart from each other about an axis ofthe tensioner spring.
 5. A tensioner as claimed in claim 1, wherein thetensioner spring is made from a wire having a generally rectangularcross-section.
 6. A tensioner as claimed in claim 1, wherein thetensioner spring is made from a wire having a generally squarecross-section.
 7. A tensioner as claimed in claim 1, wherein the dampingelement has an inside face that contacts the tensioner arm and whereinthe inside face includes channels.
 8. A tensioner as claimed in claim 1,wherein the damping element has a wall that contacts the tensioner armand wherein the wall has flex joints that extend axially along an axiallength of the wall to permit an effective reduction in radius of thewall to increase the contact area between the wall and the tensioner armunder a compressive force from the tensioner spring.
 9. A tensioner asclaimed in claim 1, wherein there is relative sliding movement betweenthe tensioner spring and the damping element and a consequent frictionaltorque provided thereby.
 10. A tensioner as claimed in claim 1, whereinthe damping member is stationary relative to the base.
 11. A drivearrangement, comprising: a crankshaft; an endless drive member driven bythe crankshaft; and a tensioner for the endless drive member, includinga shaft and a base that are mountable to be stationary relative to anengine; a tensioner arm that is pivotable relative to the shaft about atensioner arm axis; a pulley on the tensioner arm rotatable about apulley axis that is offset from the tensioner arm axis, wherein thepulley is engageable with the endless drive member; a tensioner springhaving a first end on a first end coil that is positioned to exert aforce on the shaft and the base and a second end on a second end coilthat is positioned to exert a force on the tensioner arm to urge thetensioner arm towards a free arm position; and a damping element thatengages the tensioner arm and that is engaged by a plurality of axiallyspaced segments of the tensioner spring, including a segment that is onthe second end coil having the second end, wherein the damping elementand the tensioner spring are lubricated with oil.
 12. A drivearrangement as claimed in claim 11, wherein the plurality of axiallyspaced segments are positioned on two end coils of the tensioner spring.13. A drive arrangement as claimed in claim 12, wherein the plurality ofaxially spaced segments includes at least one segment of the tensionerspring positioned between the two end coils.
 14. A drive arrangement asclaimed in claim 11, wherein the tensioner spring is generally helicaland has a first end and a second end, wherein torsional forces aretransmitted into the spring via the first and second ends, and whereinthe first and second ends are angularly 180 degrees apart from eachother about an axis of the tensioner spring.
 15. A drive arrangement asclaimed in claim 11, wherein the tensioner spring is made from a wirehaving a generally rectangular cross-section.
 16. A drive arrangement asclaimed in claim 11, wherein the tensioner spring is made from a wirehaving a generally square cross-section.
 17. A drive arrangement asclaimed in claim 11, wherein the damping element has an inside face thatcontacts the tensioner arm and wherein the inside face includeschannels.
 18. A drive arrangement as claimed in claim 11, wherein thedamping element has a wall that contacts the tensioner arm and whereinthe wall has flex joints that extend axially along an axial length ofthe wall to permit an effective reduction in radius of the wall toincrease the contact area between the wall and the tensioner arm under acompressive force from the tensioner spring.
 19. A drive arrangement asclaimed in claim 11, wherein there is relative sliding movement betweenthe tensioner spring and the damping element and a consequent frictionaltorque provided thereby.
 20. A drive arrangement as claimed in claim 11,wherein the damping member is stationary relative to the base.